1. Field of the Invention
The present invention relates to a vapor compression refrigerating apparatus (supercritical refrigerating apparatus) in which a pressure inside a gas cooler exceeds a critical pressure of a refrigerant. The present invention is applicable to a supercritical refrigerating cycle using carbon dioxide (hereinafter referred to as CO.sub.2) as a refrigerant (hereinafter referred to as CO.sub.2 cycle).
2. Description of Related Art
Theoretically, an operation of the CO.sub.2 cycle is the same as that of a conventional vapor compression refrigerating cycle using fron. That is, as indicated by line A-B-C-D-A in FIG. 24 (Mollier diagram for CO.sub.2), gas phase CO.sub.2 is compressed by a compressor (A-B), and then the gas cooler cools this high-temperature high-pressure supercritical phase CO.sub.2 (B-C).
The high-temperature high-pressure supercritical phase CO.sub.2 is decompressed by a pressure control valve (C-D) to become gas-liquid two-phase CO.sub.2. The gas-liquid two-phase CO.sub.2 is evaporated (D-A) while absorbing evaporation latent heat from external fluid such as air so that external fluid is cooled. CO.sub.2 starts phase transition from supercritical phase to gas-liquid two-phase when a pressure of CO.sub.2 becomes lower than a saturated liquid pressure (pressure at a cross point between line segment CD and saturated liquid line SL). Therefore, when CO.sub.2 performs phase transition from phase C to phase D at a slow speed, CO.sub.2 changes from supercritical phase to gas-liquid two-phase via liquid phase.
In supercritical phase, CO.sub.2 molecules move as if in gas phase even though a density of CO.sub.2 is substantially the same as that in liquid phase.
However, the critical temperature of CO.sub.2 is approximately 31.degree. C., which is lower than a critical temperature of the conventional fron (for example, 112.degree. C. for R-12). Therefore, a temperature of CO.sub.2 on a gas cooler side becomes higher than the critical temperature of CO.sub.2 during summer season or the like. Accordingly, CO.sub.2 does not condense at an outlet side of the gas cooler (line segment BC does not cross the saturated liquid line).
Furthermore, a condition of CO.sub.2 at the outlet side of the gas cooler (at point C) is determined according to a discharge pressure of the compressor and a CO.sub.2 temperature at the outlet side of the gas cooler. The temperature of CO.sub.2 at the outlet side of the gas cooler is determined by radiation performance of the gas cooler and an outside air temperature. Since the outside air temperature can not be controlled, the CO.sub.2 temperature at the outlet side of the gas cooler can not be virtually controlled.
Therefore, the condition of CO.sub.2 at the outlet side of the gas cooler (at point C) can be controlled by controlling the discharge pressure of the compressor (pressure on the gas cooler outlet side). In other words, when the outside air temperature is high during summer season or the like, the pressure of the gas cooler outlet side needs to be increased as indicated by the line E-F-G-H-E in FIG. 24, so that sufficient cooling performance (enthalpy difference) is obtained.
However, to increase the pressure on the gas cooler outlet side, the discharge pressure of the compressor has to be increased, as described above, resulting in increase in compression work (amount of enthalpy change .DELTA.L during the compression) of the compressor. Therefore, when an increasing amount of enthalpy change .DELTA.i during evaporation (D-A) is larger than an increasing amount of enthalpy change .DELTA.L during compression (A-B), a performance coefficient (COP=.DELTA.i/.DELTA.L) of the CO.sub.2 cycle deteriorates.
When calculating a relationship between the pressure of CO.sub.2 at the outlet side of the gas cooler and the performance coefficient by using FIG. 24, while setting the temperature of CO.sub.2 at the outlet side of the gas cooler to 40.degree. C., for example, the performance coefficient becomes the maximum at pressure P1 (approximately 10 MPa) as indicated by a solid line in FIG. 25. Similarly, when the temperature of CO.sub.2 at the outlet side of the gas cooler is set to 30.degree. C., the performance coefficient becomes the maximum at pressure P2 (approximately 9.0 MPa) as indicated by a broken line in FIG. 25.
Thus, each pressure in which the performance coefficient becomes the maximum is calculated for various temperatures of CO.sub.2 on the outlet side of the gas cooler in the above-mentioned method. The result is indicated by bold solid line .eta..sub.max (hereinafter referred to as optimum control line .eta..sub.max) in FIG. 24. Therefore, for an efficient operation of the CO.sub.2 cycle, the pressure on the outlet side of the gas cooler and the CO.sub.2 temperature on the outlet side of the gas cooler need to be controlled as indicated by the optimum control line .eta..sub.max.
The optimum control line .eta..sub.max is calculated so that a supercooling degree (subcooling) is approximately 3.degree. C. in a condensing area (area below the critical pressure) when the pressure on the evaporator side is approximately 3.5 MPa (corresponding to that a temperature of the evaporator is 0.degree. C.). Furthermore, FIG. 26 shows the optimum control line .eta..sub.max drawn on Cartesian coordinates having the temperature of CO.sub.2 on the gas cooler outlet side and the pressure on the gas cooler outlet side as variables. As obviously understood from FIG. 26, the pressure on the gas cooler outlet side needs to be increased as the temperature of CO.sub.2 on the gas cooler outlet side increases.
A pressure control unit for controlling a pressure on an outlet side of the gas cooler of a CO.sub.2 cycle has already been disclosed in U.S. patent application Ser. No. 08/789,210 filed Jan. 24, 1997 (corresponding Japanese patent application No. Hei 8-11248) by the inventors of the present invention et al.
In the CO.sub.2 cycle (see line A'-B'-C-D in FIG. 27), heat exchange between CO.sub.2 discharged from the evaporator (hereinafter referred to as low-pressure CO.sub.2) and CO.sub.2 discharged from the gas cooler (hereinafter referred to as high-pressure CO.sub.2) is performed so that enthalpy of CO.sub.2 at the inlet side of the evaporator is reduced, thereby increasing an enthalpy difference between the inlet and outlet sides of the evaporator to improve the cooling performance of the CO.sub.2 cycle.
However, when the inventors reviewed such CO.sub.2 cycle, it was found that the CO.sub.2 cycle may have the following problems.
In the above-mentioned CO.sub.2 cycle, the low-pressure CO.sub.2 has a preset heating degree of 0.degree. C. or more due to heat exchange between the low-pressure CO.sub.2 and the high-pressure CO.sub.2, unlike in a CO.sub.2 cycle in which heat exchange between the low-pressure CO.sub.2 and the high-pressure CO.sub.2 is not performed (see line A-B-C-D in FIG. 27).
On the other hand, the pressure control unit controls the pressure on the gas cooler outlet side according to the temperature of CO.sub.2 on the gas cooler outlet side. Therefore, the pressure control unit does not immediately reduce the pressure on the gas cooler outlet side even if the temperature of the low-pressure CO.sub.2 decreases as the heat load of the evaporator decreases and the pressure inside the evaporator decreases, but controls the pressure on the gas cooler outlet side according to the present temperature of CO.sub.2 on the gas cooler outlet side.
As a result, if the temperature of CO.sub.2 on the gas cooler outlet side does not change, the pressure on the gas cooler outlet side does not change either. Therefore, as shown in FIG. 30, when the heat load of the evaporator decreases, the temperature of CO.sub.2 increases in a CO.sub.2 passage extending from a suction side to a discharge side of the compressor. When the temperature of CO.sub.2 in the CO.sub.2 passage of the compressor is increased, shortage of oil film tends to occur at a sliding portion of the compressor, resulting in breakage of the compressor.
When the temperature of CO.sub.2 on the gas cooler inlet side increases, the temperature of CO.sub.2 on the gas cooler outlet side also increases. Therefore, when the heat load of the evaporator decreases, the pressure control unit increases the pressure on the gas cooler outlet side because the pressure control unit does not immediately respond to the temperature of the low-pressure CO.sub.2. Thus, the temperature of CO.sub.2 in the CO.sub.2 passage of the compressor may increase as the heat load of the evaporator decreases.